heat pump conceptual study and design - celeroton.com · heat pump two-phase compressor ecs...

UNCLASSIFIED

Executive summary

UNCLASSIFIED

Nationaal Lucht- en Ruimtevaartlaboratorium

National Aerospace Laboratory NLR

Report no.

NLR-CR-2014-009

Author(s)

H.J. van Gerner

Report classification

UNCLASSIFIED

Date

Knowledge area(s)

Ruimtevaarttoepassingen

Descriptor(s)

Heat Pump

two-phase

compressor

ECS

refrigeration

Heat Pump Conceptual Study and Design

Overall assessment and further work

Problem area

The classical method for spacecraft

thermal control is that the waste

heat is conducted to the radiator,

where it is radiated to space.

Therefore, the radiator temperature

must be lower than the equipment

temperature. A Heat Pump uses a

compressor to raise the radiator

temperature above the temperature

of the equipment which results in a

higher heat rejecting capacity

compared to a conventional radiator

with the same surface area.

Description of work

This report is an assessment of the

heat pump project, and contains

recommendations for further work.

Results and conclusions

In this project, a heat pump

breadboard with novel high-speed

turbo compressors has been

designed and built. The breadboard

demonstrates that the heat pump is

very efficient: At the target setting

(saturation temperature of 45°C at

the evaporator, 100°C at the

condenser, and a ‘payload’ heat

input of 5 kW), the measured COP

is 2.6, which is considerably higher

than the requirement of 2. The test

program could not be finished

completely, because particles from

the compressor sealing got stuck in

its diffusor. The recommendations

for future work include space

qualifying the compressor

electronics, increasing compressor

lifetime, and designing a receiver

that works in zero gravity.

Applicability

The heat pump developed in this

project can be used for GEO

satellites and Lunar applications.

UNCLASSIFIED

UNCLASSIFIED



Nationaal Lucht- en Ruimtevaartlaboratorium, National Aerospace Laboratory NLR

Anthony Fokkerweg 2, 1059 CM Amsterdam,

P.O. Box 90502, 1006 BM Amsterdam, The Netherlands

Telephone +31 88 511 31 13, Fax +31 88 511 32 10, Web site: www.nlr.nl

http://www.nlr.nl/

Nationaal Lucht- en Ruimtevaartlaboratorium

National Aerospace Laboratory NLR

NLR-CR-2014-009



H.J. van Gerner

No part of this report may be reproduced and/or disclosed, in any form or by any means without the prior

written permission of the owner.

Customer European Space Agency

Contract number ESTEC Contract No.4000105360/12/NL/EM

Owner NLR + partner(s)

Division NLR Aerospace Systems and Applications

Distribution Limited

Classification of title Unclassified

Approved by:

Author

Henk Jan van Gerner

Reviewer

Johannes van Es

Managing department

Michel Keuning

Date: Date: Date:

NLR-CR-2014-009

3

Summary

The classical method for spacecraft thermal control is that the waste heat is conducted to the

radiator where it is radiated to space. Therefore, the radiator temperature must be lower than the

equipment temperature. A Heat Pump uses a compressor to raise the radiator temperature above

the temperature of the equipment which results in a higher heat rejecting capacity compared to a

conventional radiator with the same surface area. For example, when a heat pump is used to lift

the radiator temperature from 45°C to 100°C, the heat rejecting capacity of a GEO satellite can

be increased with 83%. However, commercially available compressors have a high mass (40 kg

for 10kW cool capacity), cause vibrations, and are intended for much lower temperatures

(maximum 65°C) than what is required for the heat pump application (100°C). Dedicated

aerospace compressors have been developed with a lower mass (19 kg) and for higher

temperatures, but these compressors have a lower efficiency.

In the preliminary design report for this project, a heat pump system was presented with a novel

electrically-driven turbo compressor system with a mass of 2kg and a higher efficiency than

existing aerospace compressors. The selected fluid for the heat pump is isopentane. A detailed

design for a heat pump breadboard was presented in the detailed design report for this project.

The assembly of the breadboard, the test plan and procedures are discussed in the test readiness

report, and the test results are discussed in the breadboard test report. This report is an

assessment of the heat pump project, and contains recommendations for further work.

At the target setting (saturation temperature of 45°C at the evaporator, 100°C at the condenser,

and a ‘payload’ heat input of 5 kW), the measured COP is 2.6, which is considerably higher

than the requirement of 2. The COP is higher because the efficiency of the compressors is

higher than expected, and the pressure drop in the evaporator is lower. During one of the tests,

one of the compressors started to make a noise and showed a decrease in performance. After

opening the compressor, it was found that particles from the inlet and outlet connection sealing

got stuck in the diffusor of a compressor. As a result, the test program could not be finished

completely.

The recommendations for future work include space qualifying the compressor electronics,

increasing the lifetime of the compressor, a method to cool the compressor with its own fluid,

and designing a receiver that works in zero gravity.

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Contents

1 Introduction 6

1.1 Background 6

1.2 Heat pump working principles 6

1.3 Heat pump breadboard characteristics and fluid properties 8

2 Assessment of the results 10

2.1 Heat pump breadboard 10

2.2 Heat pump compressors 14

2.2.1 General design 14

2.2.2 Comparison with existing compressors 14

2.3 Measured COP and compressor efficiency 15

2.4 Noise in the Stage 3 compressor and subsequent investigation results 16

2.5 Summary for the test program 17

3 Recommendations for future work 19

3.1 Space qualify compressor electronics 19

3.2 Increasing the life time of the compressors 19

3.3 Cooling of the compressors with its own fluid 19

3.4 Using passive expansion valves instead of electronic expansion valves 19

3.5 Designing a receiver that works in zero gravity 20

3.6 Increasing the COP of the system 20

3.6.1 Increasing the compressor efficiency 20

3.6.2 Reduce the pressure drop in the system by using parallel evaporators 20

3.6.3 Use additional heat exchangers and/or valves to increase the COP 21

3.7 Use a Liquid-Suction Heat Exchanger (LSHE) to obtain the required superheat

for the compressors 22

3.8 Make a full size breadboard 22

3.9 Perform environmental (e.g. vibration and thermal cycling) tests 22

4 Conclusions 23

References 24

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Applicable documents 24

Appendix A Increasing the COP with a Liquid-Suction Heat Exchanger 25

Appendix B Increasing the COP with flash economizers 27

Appendix C Increasing the COP with heat exchanger economizers 29

Appendix D Increasing the COP with cascades 31

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1 Introduction

1.1 Background

The classical method for spacecraft thermal control is that the waste heat is conducted to the

radiator, where it is radiated to space. Therefore, the radiator temperature must be lower than the

equipment temperature. In addition, the amount of heat rejected by a radiator is directly driven

by the surface temperature and surface area. In essence, the limiting factors of the radiator

performance are the equipment operating temperature and launch volume. Therefore, in order to

increase the heat rejection capability using standard heat transfer method, a deployable radiator

can be used to increase the radiating surface area or the operating temperature of the internal

units can be increased. Another method is to increase the radiator temperature without

increasing the temperature of the internal components. This is the major characteristic of a Heat

Pump which makes this product unique since it has the capability of raising the radiator

temperature above the temperature of the equipment where more heat can be rejected compared

to a conventional radiator with the same surface area.

However, this increase in heat rejection capacity comes at price of a large mass penalty;

Commercial available compressors with a cooling capacity of 10 kW at a ΔT of 55 K have a

very large mass (e.g. 40 kg or 80 kg when for redundancy two compressors are used).

Furthermore, the extra mass caused by the electric power for the compressor (due to larger solar

panels, battery, and power control unit) is approximately 125 kg (assuming 25 kg/kW). For this

reason, it is of utmost importance to use a very lightweight compressor with a high Coefficient

of Performance (COP).

These characteristics are combined in the miniature electrically-driven compressors made by

Celeroton. These compressors run at ultra-high speeds of up to 500,000 RPM, which results in

very low masses, approximately 2 kg for a cooling capacity of 10 kW.

In the preliminary design report for this project [1], a heat pump system was presented with a

novel electrically-driven turbo compressor system with a mass of 2kg and a higher efficiency

than existing aerospace compressors [5-7]. The selected fluid for the heat pump is isopentane. A

detailed design for a heat pump breadboard was presented in the detailed design report for this

project [2], and a review of the assembled breadboard was presented in the test readiness report

[3]. The test results of the breadboard are discussed in [4]. This report gives an assessment of

the project and gives recommendations for future work.

1.2 Heat pump working principles

A vapor compression Heat Pump consists of a compressor, a heat exchanger at the hot source

(i.e. the evaporator), a heat exchanger at the cold sink (i.e. the condenser), and an expansion

NLR-CR-2014-009

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valve (see Fig. 1 for a schematic drawing). The working fluid enters the compressor as a

superheated vapour. The compressor increases the pressure and the temperature of the vapour.

The vapour then travels through the condenser, where the vapour is condensed into liquid and

the heat that is stored in the vapour is released. The liquid flows through the expansion valve,

where the pressure abruptly decreases (adiabatic expansion), causing a partial evaporation of the

liquid and a drop in the temperature. The cold liquid-vapour mixture then flows through the

evaporator where it absorbs heat and completely turns into vapour before entering the

compressor.

Fig. 1 Schematic drawing of a basic vapour compression cycle

The heat pump cycle with isopentane (R601a) as working fluid and where the saturation

temperature is increased from 45 to 100°C is represented in the enthalpy-pressure diagram in

Fig. 2. In the ideal cycle (solid line), the fluid leaves the evaporator and enters the compressor as

saturated vapour. In an actual vapour compression cycle (dashed line), the vapour is slightly

superheated (to 50°C) to ensure that the fluid is completely vaporized before it enters the

compressor. In the ideal cycle, the compression process is isentropic, while in an actual cycle,

the adiabatic efficiency of the compression process is approximately 60%. Furthermore, there is

a pressure drop (which is assumed to be 0.4 bar) in the condenser, evaporator and transport lines

of an actual heat pump.

evaporator

condenser

recie

ve

r

inQ

outQ

vapour

+ liquid

vapour

TXV

liquid

liquid vapour

Tsat=45°C

Tsat=100°C

2 3

4 1

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Fig. 2 Pressure-Enthalpy diagram of a vapour compression cycle with isopentane (R601a) as refrigerant. The ‘skewed dome’ in the diagram is the two-phase region, i.e. the fluid in that region is a mixture of liquid and vapour. To the right of the dome, the fluid is vapour, to the left, the fluid is liquid. The green lines in the diagram are isentropic lines, the blue lines are isodensity lines, and the dark red lines are isothermal lines.

1.3 Heat pump breadboard characteristics and fluid properties

The main characteristics of the heat pump breadboard are summarized in Table 1. Some

properties of the selected working fluid are listed in Table 2.

Saturation temperature in the evaporator 45°C

Saturation temperature in the condenser 100°C

Heat pump cooling capacity

(evaporator heat load) 5 kW

Total compressors power ~2.5 kW divided over 3 compressors in a

serial configuration

Condenser capacity 7.5 kW (5 kW for heat load and 2.5 kW

for compressor power)

Target COP 2

Working fluid Isopentane (R601a)

Internal volume of system ~3 liters total (~1 liter without the

accumulator and receiver)

evaporation

condensation

expansion

effect of pressure drop

superheat

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Table 1 Heat pump breadboard characteristics [1]

Saturation pressure at 45°C 1.76 bar

Saturation pressure at 100°C 7.20 bar

Triple point temperature -161°C

Critical temperature 187°C

Heat of evaporation hlv at 45°C 0.33 MJ/kg

Vapour density ρv at 45°C 5.14 kg/m3

Liquid density ρl at 20°C 620 kg/m3



Thermal conductivity k at 45°C 0.10 W/(m K)

Table 2 Fluid properties of isopentane (R601a). The fluid properties are taken from REFPROP [8]

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2 Assessment of the results

2.1 Heat pump breadboard

In this project, a heat pump breadboard with three novel electrically-driven high-speed turbo

compressors in a serial configuration has been designed and built. The refrigerant for the

breadboard is isopentane (R601a). Fig. 3 shows a schematic layout of the heat pump

breadboard. Isopentane enters the first compressor as a slightly superheated (to 50°C) vapour.

Three compressors are used in a serial configuration to increase the saturation temperature of

the fluid to the desired value (100°C) [1]. Before and after each compressor, the pressure and

temperature is measured, so that the isentropic efficiency of each compressor can be calculated,

as well as the efficiency of the total compressor system. After the compressors, vapour travels

through the condenser, where the vapour is condensed into liquid and the heat that is stored in

the vapour is released. The temperature of the condenser is controlled with a thermostat bath

and cooling water. The liquid then flows into the receiver at the top, and out of the receiver at

the bottom. After the receiver, the liquid flows through the expansion valve, where the pressure

abruptly decreases, causing a partial evaporation of the liquid (to a vapour mass fraction of 0.36,

see Fig. 2) and a drop in the temperature. The cold liquid-vapour mixture then flows through the

evaporator where it absorbs heat and completely turns into vapour before entering the

compressor.

A CAD drawing of assembled system is shown in Fig. 4 and a photo is shown in Fig. 5. A

detailed description of the breadboard including a list of all the components is provided in [2].

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Fig. 3 Schematic drawing of the heat pump breadboard

condenser

(plate heat

exchanger)

receiver

evaporator

(plate heat

exchanger)

T1 T2P1

P2

Exhaust / purge

Dout=15 mm

DA

DA Data Acquisition

FM Flow Meter

PF Particle Filter

PID proportional–integral–derivative controller

PS Power Supply

P Pressure Transducer

TXV Thermal Expansion Valve

T Temperature Transducer

V Valve

TXV

PID2

0

0

0

0

0

T6

T7

T3

P5

Do

ut=

15

mm

V4

accumulator

thermostatic

bath at 80 °C

PF1

PF2

thermostatic

bath at 20 °C

PID1

0

0

0

0

0

P3

P4

V1

V2

vacuum pump

V3

supply

PS3

(230VAC)

u1x1

f(x1...xn)

T4

T5

Dout=12mm

Do

ut=

12

mm

FM1

Dout=15mm

burst disk (12 bar)

burst disk

(8 bar)

T24

T23

T22

T21

FM2

T8

thermostatic

bath at 20 °C

thermostatic

bath at 60°C

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Fig. 4 CAD drawing of the assembled system.

flowmeter

evaporator

condenser

compressors

expansion

valve

receiver

accumulator

Pressure sensors

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Fig. 5 Photo of the breadboard.

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2.2 Heat pump compressors

2.2.1 General design

Three electrical-driven turbo compressors made by Celeroton are used in a serial configuration.

For the first stage, an (slightly modified) existing CT-17-700 turbo compressor with a 3D

impeller (see Fig. 6) has been used. For the second and third stage, a new turbo compressor with

a 2D impeller (see Fig. 7) has been designed. The compressors run approximately at 180 kRPM.

The design and manufacturing of the compressor is described more extensively in [6-8].

Fig. 6: Drawing of the 3D-impeller for the first stage

Fig. 7: Drawing of a 2D-impeller for the second and third stage

2.2.2 Comparison with existing compressors

In [1], a comparison has been made between commercial and aerospace compressors, and the

compressors developed in this project. The results of the compressor comparison are

summarized in Table 3. From this table, it can be concluded that the compressor developed in

this project have:

A mass of 2 kg, compared to >20 kg for other compressors with the same capacity

A vibration level of <0.3 N compared to 40 N for commercial compressors

An isentropic efficiency of 68%, which is higher than aerospace compressors (46%), and

comparable to commercial compressors (60-70%)

A condenser saturation temperature of 100°C, which is comparable to aerospace

compressors, but higher than the maximum temperature of commercial compressors (65°C)

No oil required for lubrication or sealing, compared to commercial compressors which

require a few litres of oil.

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Compressor Refrigerant

Mass

for 10

kW cool

capacity efficiency

Max

condenser

temperature

Requires

oil for

lubrication

and sealing

Estimated

vibration

level

Copeland

ZB30KCE R134a 40 kg 59.2% 65°C

Yes, 1.9

litres ~40N

Copeland

ZB42KCE R134a 40 kg 66.3% 65°C

Yes, 1.9

litres ~40N

Danfoss

MLZ045 R22 37 kg 67.8%. 45°C

Yes, 1.6

litres ~40N

Danfoss

MLZ045 R404a 37 kg 70.3% 45°C

Yes, 1.6

litres ~40N

Danfoss The

MLZ076 R134a 45 kg 59.6% 45°C

Yes, 2.7

litres ~40N

Fairchild 54

mm Twin

Screw

? 19 kg <50%?? 93.3°C Yes ?

Sundstrand

LANTIRN

compressor

R114

10 kg

for 2.5

kW

Max

45.7% 110°C Yes >> 0.3N

Honeywell

F-22

compressor

? ? ? ? ? ?

Compressor

developed in

this project

Isopentane

(R601a) ~2 kg 68% 100°C No < 0.3N

Table 3 Comparison table

2.3 Measured COP and compressor efficiency

Fig. 8 shows the measured heat pump cycle in a pressure-enthalpy diagram. At the target setting

(saturation temperature of 45°C at the evaporator, 100°C at the condenser, and a ‘payload’ heat

input of 5 kW), the measured COP is 2.6, which is considerably higher than the requirement of

2. This is because the pressure drop in the evaporator is lower (0.2 bar instead of 0.4 bar), and

because the efficiency of the compressors is higher than expected (0.68 instead of 0.6). The

measured efficiency of 0.68 of the compressors is considerable higher than existing aerospace

compressors, and even higher than the efficiency of most of the commercially available scroll

compressors that were analysed in the compressor comparison (see previous section).

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Fig. 8 Measured (red) and target (blue) heat pump cycle in a pressure-enthalpy diagram. The measured cycle has a COP of 2.6, which is higher than the requirement of 2

2.4 Noise in the Stage 3 compressor and subsequent investigation results

During a measurement, a loud noise was observed. Furthermore, the efficiency of the stage 3

compressor became lower than in previous measurements. For this reason, it was concluded that

the stage 3 compressors could not be used anymore, and it was send back to Celeroton for

inspection. The analysis at Celeroton showed that large particles are stuck in the diffusor (see

Fig. 9 and Fig. 10 for a detail picture of the particles that were stuck in the diffusor). This

resulted in blockage of some flow channels, which led to a surge at higher mass flows which

probable resulted in the observed noise and vibrations. Also it is obvious that this resulted in a

lower efficiency. The most likely origin of the particles are the Viton sealing rings that are used

to seal the inlet and outlet connections of the compressors. In a subsequent project, a different

sealing method for the inlet and outlet connections must be used. This is a relative simple

modification.

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Fig. 9 Photo of the rotor and diffusor of the stage 3 compressor, where it can be seen that large particles are stuck in the diffusor

Fig. 10 Detail picture of the particles that were stuck in the

diffusor

2.5 Summary for the test program

Because of the blockage of the stage 3 compressor with particles from the seals of the

compressor (see previous section), only 5 out 8 items of the test plan have been carried out. The

maximum cooling capacity test, short life time test, and the increase in cooling water

temperature test have not been carried out. For the maximum cooling capacity test, it is assumed

that the cooling capacity of 5 kW could be increased somewhat, because the cooling capacity of

5 kW was achieved with 175 kRPM, while the maximum RPM of the compressors is set to 200

kRPM. The compressor cooling water temperature used in the test was 25°C. The motor

temperature is 80°C. Since the motors are allowed to have a temperature of 100°C, it is

expected that the thermostat water temperature could be increased to 45°C. A summary of the

test program is shown in Table 4.

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Test description Test results remark

Proof pressure test successful After modification of the

compressors

Helium leak test successful The system has been helium

leak tested 5 times

Filling of the breadboard with

isopentane successful

The system has been filled

with isopentane 5 times

Start-up test successful

The system has been started-

up around 25 times. After

some modifications to the

compressor and condenser

settings, the start-up is

straight forward

COP test successful, 2.6 measured

The measured COP of 2.6, is

significantly higher than the

requirement of 2

Max cooling capacity test not carried out

It is expected that the cooling

capacity can be increased

above 5 kW

Short life time test not carried out

Increase in cooling water

temperature test not carried out

It is expected that the cooling

water temperature can be

increased to 45°C

Table 4 Summary table for the tests that has been carried out

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3 Recommendations for future work

3.1 Space qualify compressor electronics

The converters that drive the compressors are not space qualified, not optimized for smallest

mass and size, and not adapted to the final dc voltage bus of spacecraft. With a specific

electronics design, all these drawbacks could be resolved. Celeroton directly has no experience

in designing space qualified electronics, however the board of directors has contacts with

companies which themselves design space qualified electronics and offer consulting in space

qualified electronics. Furthermore, NLR has experience in designing and qualifying electronics

for space applications.

3.2 Increasing the life time of the compressors

The compressor for this project phase have been equipped with ball bearings, as the goal of this

project was to prove the thermodynamic feasibility. Therefore, the lifetime is not yet sufficient

for space missions. Therefore, one main proposed future task is the replacement of the ball

bearings with a fluid or magnetic bearing, which are both contactless, and therefore show no

mechanical wear and a long lifetime. There are several technical challenges to solve such as

fixation or levitation during rocket start (vibrations).

Since it was observed during the test phase of the heat pump breadboard that particles from the

inlet and outlet connection sealing got stuck in the diffusor of a compressor, it is necessary to

redesign the inlet and outlet sealing of the compressors. However, this is relative simple.

3.3 Cooling of the compressors with its own fluid

In the current breadboard, the motors of the compressors are cooled with water from a

thermostat bath. It is proposed to cool the motors with isopentane from the heat pump. For this,

in a first step the water of the thermostat bath can be increased to 45°C, and in a second step the

cooling loop can be changed to isopentane. The best solution is to use ispentane coming from

the expansion valve, since this fluid has a low temperature (45°C) and the pressure drop induced

by the motor cooling circuit has no negative influence on the performance of the heat pump.

3.4 Using passive expansion valves instead of electronic expansion valves

The expansion valve regulates the superheat (e.g. 5°C) after the evaporator: If the superheat is

too large, the valve opens more which results in a larger massflow and thus smaller superheat.

Vice versa, when the superheat is too small, the valve closes more which results in more

superheat. There are two basic types of expansion valves: Electronic expansion valves and

thermostatic expansion valves. In the current breadboard, an electronic expansion valve is used

NLR-CR-2014-009

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to control the superheat temperature. In an electric expansion valve, the valve opening is

adjusted by an electric actuator via a PID controller that regulates the superheat temperature to

the desired value. An electronic expansion valve can therefore be used for a wide range of

fluids. However, the electronics required for an electronic expansion valve are undesired for a

space application.

A thermostatic expansion valve regulates this superheat by use of a temperature sensing bulb

filled with the same fluid as in the system. Thermal expansion valves are not available for every

fluid, because the sensing bulb must be filled with the same (or similar) fluid as in the system.

For the most common refrigerants (R22, R134a, R407c, R404a, ammonia), these valves are

easily available, but there are no thermostatic expansion valves available for isopentane.

Danfoss also makes thermostatic expansion valves for hydrocarbons, but not for isopentane.

However, it is fairly easy to adapt an existing thermal expansion valve to isopentane and for this

reason, it is proposed to use a thermostatic expansion valve in a future breadboard.

3.5 Designing a receiver that works in zero gravity

The receiver in the breadboard depends on gravity to ensure that only liquid is always present at

the exit of the receiver. It is proposed to design a receiver that also functions in zero gravity

(and preferably in all orientations), for example by using centrifugal forces.

3.6 Increasing the COP of the system

There are several methods for increasing the COP of a heat pump. These are discussed in the

next sections.

3.6.1 Increasing the compressor efficiency

An obvious method to increase the COP of a system is by increasing the efficiency of the

compressors. The measured efficiency of 0.68 is already very high, but there are measures that

could allow for an even further increase of the efficiency.

Compressor stage 2 and 3 are realized as pure 2D impellers. These types of Impellers show a

high influence of the tip clearance on the pressure ratio and efficiency. This is known from

theory and could also be shown during first tests in air by adjusting the tip clearance. For such

miniature, ultra-high-speed compressors, the tip clearance has a lower limit due to tolerances

and vibrations. A shroud could lower the influence in the efficiency drop and in the required

tolerances. Therefore, it is proposed to include also shrouded impellers in future analysis and

redesigns.

3.6.2 Reduce the pressure drop in the system by using parallel evaporators

The pressure drop in the system, and especially the pressure drop in the evaporator, has a very

large influence on the COP of the system. This is illustrated by Fig. 11, which shows the

NLR-CR-2014-009

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calculated COP of the heat pump, as a function of the pressure drop over the evaporator and

condenser. In this calculation, the efficiency of the compressors is assumed to be 68%.

In the experiments, the pressure drop is 0.2 bar, and the COP is 2.6. However, in an actual

application, the evaporator can exist of long tubing, and the pressure drop can become much

higher. There are several methods to reduce the pressure drop in a system. The most

straightforward method is to increase the diameter of the tubing. However, this also increases

the volume of the system, which is undesired. Another method is to use parallel evaporator

sections. With parallel channels, the diameter of the tubing can remain small, and it is therefore

proposed to test parallel evaporators in a future breadboard.

Fig. 11 COP of the system as a function of the pressure drop over the evaporator and condenser. The efficiency of the compressors is assumed to be 68%.

3.6.3 Use additional heat exchangers and/or valves to increase the COP

For a ‘wet’ refrigerant like isopentane, the two-phase dome is ‘tilted’ to the right (see Fig. 2).

As a result, the liquid/vapour mixture that enters the evaporator has a relative high vapour mass

fraction of almost 0.4. This means that a large portion of the latent heat of the fluid is not used

in the evaporator. If the vapour mass fraction of the liquid/vapour mixture that enters the

evaporator is reduced to for example 0.1, a much larger portion of the latent heat of the fluid is

used in the evaporator, which reduces the massflow in the system and therefor increases the

COP of the cycle. There several methods to achieve this:

Using a Liquid-Suction Heat Exchanger (see Appendix A)

Using Flash Economizers (Appendix B)

Using Heat Exchanger Economizers (Appendix C)

Using three cascaded cycles (Appendix D)

In Appendix A to D, these methods are analysed for the heat pump cycle with isopentane. With

all four methods, the COP of the system can be increased by 17 to 21%. A Liquid-Suction Heat

NLR-CR-2014-009

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Exchanger (LSHE) is the most simple method, and requires just one additional heat exchanger

in the system. However, in practice there will be a pressure drop over the heat exchanger (this

pressure drop is not included in the analysis, since it depends on the design of the heat

exchanger), and this pressure drop can have a negative influence on the COP of the system (see

Fig. 11). A system with Heat Exchange Economizers is more complex, but the vapour massflow

through the heat exchangers is very small (just 14% of the total massflow) and as a result, the

pressure drop over the heat exchanger will be very small.

The method with Flash Economizers requires two additional expansion valves and two

separation vessels, which separate the liquid from the vapour. For a space application,

centrifugal forces in the vessel can be used to separate the vapour from the liquid, and this could

be a very lightweight solution.

It is proposed to make a trade-off between the different methods to increase the COP by adding

heat exchangers and/or valves

3.7 Use a Liquid-Suction Heat Exchanger (LSHE) to obtain the required superheat for the

compressors

It was observed in the measurements that due to the high efficiency of the compressors, the

required superheat of the vapour that enters the compressors has to be rather larger. This

superheat is inherent provided when a Liquid-Suction Heat Exchanger (see Appendix A) is

used. An additional advantage of a LSHE is that it increases the COP of the heat pump. A

LSHE does not provide superheating for the compressors during start-up, so this is an issue that

has to be addressed.

3.8 Make a full size breadboard

In the current breadboard, the evaporator and condenser that are used are not representative for a

space application concerning volume, tubing length etc. This can have influences on the time

dependent behaviour (e.g. during start-up) of the system. It is therefore recommended to make a

full size breadboard.

3.9 Perform environmental (e.g. vibration and thermal cycling) tests

For a space application, a heat pump is subjected to severe vibrations (during launch) and

temperature variation. It is there recommended to perform relevant environmental tests on the

complete system or in an earlier phase on the most critical components (e.g. the compressors,

electronics, and expansion valve).

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23

4 Conclusions

In this project, a heat pump breadboard with three novel electrically-driven high-speed turbo

compressors in a serial configuration has been designed and built. The refrigerant for the

breadboard is isopentane (R601a). The breadboard demonstrates that the used heat pump

method is very efficient: At the target setting (saturation temperature of 45°C at the evaporator,

100°C at the condenser, and a ‘payload’ heat input of 5 kW), the measured COP is 2.6, which is

considerably higher than the requirement of 2. The compressors developed in this project have a

much lower mass and higher efficiency than existing (commercial or aerospace) compressors.

The test program could not be finished completely, because particles from the sealing of the

compressor got stuck in its diffusor

The recommendations for future work include space qualifying the compressor electronics,

increasing the lifetime of the compressor, a method to cool the compressor with its own fluid,

and designing a receiver that works in zero gravity.

NLR-CR-2014-009

24

References

[1] H.J. van Gerner, Heat Pump Conceptual Study and Design: Preliminary Design Report,

NLR-CR-2012-177 (2012)

[2] H.J. van Gerner, Heat Pump Conceptual Study and Design: Detailed Design Report,

NLR-CR-2012-424 (2013)

[3] H.J. van Gerner, A. Pauw, G. van Donk, Heat Pump Conceptual Study and Design: Test

Readiness Report, NLR-CR-2013-327 (2013)

[4] H.J. van Gerner, A. Pauw, G. van Donk, Heat Pump Conceptual Study and Design:

Breadboard Test Report, NLR-CR-2014-007 (2014)

[5] Celeroton AG, Heat Pump Conceptual Study and Design: WP 4100 Turbocompressor

Preliminary Design (2012)

[6] Celeroton AG, Heat Pump Conceptual Study and Design: WP 4200 Compressor

Detailed Design, PR-6901-001 (2013)

[7] Celeroton AG, Heat Pump Conceptual Study and Design: WP 4300 Turbocompressor

Manufacturing, PR-6901-002 (2013)

[8] Lemmon, E.W., Huber, M.L., McLinden, M.O. NIST Standard Reference Database 23:

Reference Fluid Thermodynamic and Transport Properties-REFPROP, Version 8.0,

National Institute of Standards and Technology, Standard Reference Data Program,

Gaithersburg, 2007.

Applicable documents

[AD1] Statement of Work for ESA Invitation To Tender AO/1-6751/11/NL/EM

[AD2] NLR proposal SID6393, Conceptual heat pump study and design, issue 2, dated

January 24th 2012.

[AD3] H.J. van Gerner, Heat Pump Conceptual Study and Design: Preliminary Design

Report, NLR-CR-2012-177 (2012)

[AD4] H.J. van Gerner, Heat Pump Conceptual Study and Design: Detailed Design

Report, NLR-CR-2012-424 (2013)

[AD5] H.J. van Gerner, A. Pauw, G. van Donk, Heat Pump Conceptual Study and

Design: Test Readiness Report, NLR-CR-2013-327 (2013)

[AD6] Celeroton AG, Heat Pump Conceptual Study and Design: WP 4100

Turbocompressor Preliminary Design (2012)


Compressor Detailed Design, PR-6901-001 (2013)


Turbocompressor Manufacturing, PR-6901-002 (2013)

NLR-CR-2014-009

25

Appendix A Increasing the COP with a Liquid-Suction Heat Exchanger

For a ‘wet’ refrigerant like isopentane, the two-phase dome is ‘tilted’ to the right. As a result,

the liquid/vapour mixture that enters the evaporator has a relative high vapour mass fraction of

almost 0.4. This means that a large portion of the latent heat of the fluid is not used in the

evaporator. When the liquid coming out of the condenser is subcooled to 60°C, the vapour mass

fraction of the liquid/vapour mixture that enters the evaporator is reduced to 0.1 (see Fig. 13).

As a result, a lower mass flow (and therefor lower compressor power) is required to achieve the

same cooling capacity, and this increases the COP. A Liquid-Suction Heat Exchanger (LSHE)

uses the cold vapour coming from the evaporator to cool the hot liquid from the condenser (see

Fig. 12 and Fig. 13). As a result, the vapour that enters the compressor is at a higher temperature

(92°C for the cycle in Fig. 13). The exit temperature of the compressor therefore also increases

from 115°C without LSHE to 156°C with LSHE. The COP is increased from 2.04 without

LSHE, to 2.41 with LSHE (i.e. an 18% increase in the COP). The drawbacks of a LSHE are the

additional mass of the heat exchanger, and the higher compressor exit temperature.

Fig. 12 Schematic drawing of a vapour compression cycle with a Liquid Suction Heat Exchanger

eva

po

rato

r

co

nd

en

se

r

inQ outQ

vapour

liquid

vapour

TXV

he

at e

xch

an

ge

r

liquid vapour

1 2 3 4

5

6

7

8

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Fig. 13 Vapour compression cycle with a Liquid-Suction Heat Exchanger. The refrigerant is isopentane

Liquid-Suction

Heat Exchange

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Appendix B Increasing the COP with flash economizers




evaporator. The vapour mass fraction of the fluid that enters the evaporator can be reduced by

not expanding the fluid with one valve, but by using multiple valves (see Fig. 14). After the first

expansion valve, the mixed vapour/liquid flow is separated into vapour and liquid. For

terrestrial applications, this is usually achieved in a tank due to gravity. However, for a space

application, centrifugal forces can be used to separate the vapour from the liquid. The vapour

(which is 14% of the total mass flow for the cycle depicted in Fig. 15) is then injected between

the second and third compressor stages. The liquid (86% of the total mass flow) is directed to a

second expansion valve. After the second expansion valve, the mixed vapour/liquid flow is

again separated into vapour and liquid, and the vapour is injected between the thirst and second

compressor stage. The liquid is directed to the third expansion valve. The vapour mass fraction

of the liquid/vapour mixture that enters the evaporator is reduced to 0.1 (compared to 0.4 for a

basic cycle, see Fig. 2). This means that a much larger portion of the latent heat of the fluid is

used in the evaporator, which increases the COP of the cycle. For the cycle with isopentane, the

COP is increased from 2.03 without flash economizers, to 2.46 with flash economizers (i.e. a

21% increase in the COP).

Fig. 14 Schematic drawing of a vapour compression cycle with a Flash Economizer

eva

po

rato

r

co

nd

en

se

r

inQ outQ

vapour

TXV1

liquid

vapour

TXV3

liquid

(~73% of massflow)

vapour

(~14% of massflow)

TXV2

vapour

(~13% of massflow)

liquid

(~86% of massflow)

1 2 3 4 5 6

7

8

9

10

11

12

14

13

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Fig. 15 Vapour compression cycle with a Flash Economizers.

NLR-CR-2014-009

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Appendix C Increasing the COP with heat exchanger economizers




evaporator. The vapour mass fraction of the fluid that enters the evaporator can be reduced by

not expanding the fluid with one valve, but by using multiple valves and heat exchangers (see

Fig. 16). Before the first expansion valve, the liquid flow is divided into a main flow (86% of

the massflow for the cycle depicted in Fig. 17) and a secondary flow (14% of the massflow). In

a heat exchanger, the main flow is cooled by the secondary flow. When the secondary flow

absorbs heat from the primary flow, it evaporates. The secondary vapour flow is then injected

between the second and third compressor stage. Before the second valve, the flow is again

separated, and the main flow is further cooled to 62°C before it enters the third expansions

valve. As a result of the stepwise cooling of the main flow, the vapour mass fraction of the

liquid/vapour mixture that enters the evaporator is reduced to 0.1 (compared to 0.4 for a basic

cycle, see Fig. 2). This means that a much larger portion of the latent heat of the fluid is used in

the evaporator, which increases the COP of the cycle. For the cycle with isopentane, the COP is

increased from 2.03 without heat exchange economizers, to 2.39 with heat exchange

economizers (i.e. a 17% increase in the COP). Because the vapour flow through the heat

exchangers is very small (~14% of the total flow), the pressure drop over the heat exchanger

will be very small.

Fig. 16 Schematic drawing of a vapour compression cycle with a Heat Exchanger Economizer

eva

po

rato

r

co

nd

en

se

r

inQ outQ

vapour

TXV1

vapour

TXV3

liquid

(~86% of

massflow)

he

at e

xch

an

ge

r

TXV2

he

at e

xch

an

ge

r

1 2 3 4 5 6

7 8 9 10 11 12

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Fig. 17 Vapour compression cycle with a Heat Exchanger Economizer.

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Appendix D Increasing the COP with cascades

The heat pump cycle can be separated in three cycles (see Fig. 18). Cycle b is thermally

connected to cycle a, and cycle b is thermally connected with cycle c. With three cascades, the

vapour mass fraction of the liquid/vapour mixture that enters the evaporator is reduced to 0.1

(compared to 0.4 for a basic cycle, see Fig. 2). This means that a much larger portion of the

latent heat of the fluid is used in the evaporator, which increases the COP of the cycle. For the

cycle with isopentane (see Fig. 19), the COP is increased from 2.03 with one cycle, to 2.47 with

3 cascaded cycles (i.e. a 21% increase in the COP). In Fig. 19, isopentane is used in all three

cycles. However, it is also possible to use different refrigerants for each cycle.

Fig. 18 Schematic drawing of a vapour compression cycle in three cascades

Fig. 19 Vapour compression cycle in three cascades.

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po

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co

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en

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inQ outQ

TXV1TXV3

he

at e

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an

ge

r

TXV2

he

at e

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an

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r

1a 2a

3a 4a

1c 2c

3c 4c

1b 2b

3b 4b

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